Centrifugal pump

ABSTRACT

A centrifugal pump for conveying a process fluid, includes a pump unit, a drive unit to drive the pump unit, a pump inlet to receive the process fluid, a pump outlet to discharge the process fluid. The pump unit includes an impeller configured to convey the process fluid from the pump inlet to the pump outlet, and a pump shaft, on which the impeller is mounted. The drive unit includes a drive shaft to drive the pump shaft, and an electric motor to rotate the drive shaft about an axial direction. A plurality of bearings is configured to support the pump shaft and the drive shaft, and a hydrodynamic coupling has a casing and is configured to hydrodynamically couple the drive shaft to the pump shaft by a transmission fluid. At least one of the plurality of bearings is arranged in the casing of the hydrodynamic coupling.

CROSS-REFERENCE TO RELATED APPLICATION

This application claims priority to European Patent Application No.19170572.2 filed Apr. 23, 2019, the content of which is herebyincorporated herein by reference.

BACKGROUND Field of the Invention

The invention relates to a centrifugal pump for conveying a processfluid.

Background Information

Conventional centrifugal pumps have at least one rotating impeller forconveying a process fluid from a pump inlet to a pump outlet. The atleast one impeller can be configured as a radial, axial, mixed flow orhelico-axial impeller. Centrifugal pumps are known in a great variety ofembodiments such as single stage pumps, multistage pumps, single phasepumps or multiphase pumps, just to list a few examples. Centrifugalpumps are used in many different industries, for example in the cleanand waste water industry, in the chemical processing industry or thepower generation industry. Another important example is the oil and gasprocessing industry, where centrifugal pumps are designed, e.g. asmultiphase pumps for conveying hydrocarbon fluids, for example forextracting the crude oil from the oil field or for transportation of theoil/gas through pipelines or within refineries. Another application ofcentrifugal pumps in the oil and gas industry is the injection of aprocess fluid, in most cases water and in particular seawater, into anoil reservoir. For such applications, the pumps are designed as waterinjection pumps for supplying seawater at high pressure to a well thatleads to a subterranean region of an oil reservoir. A typical value forthe pressure increase generated by such a water injection pump is200-300 bar (20-30 MPa) or even more. For such applications it is alsoknown to configure the centrifugal pump as a process fluid lubricatedpump, i.e. a pump which uses the process fluid as lubricant and/orcoolant for the bearings supporting the rotating shaft. In addition, ina process fluid lubricated pump the process fluid can also be used forcooling the motor that drives the rotation of the pump shaft.

Water injection into oil reservoirs is a well-known method forincreasing the recovery of hydrocarbons from an oil or gas field. Theinjected water maintains or increases the pressure in the reservoirthereby driving the oil or the hydrocarbons towards and out of theproduction well.

In some applications, raw seawater is injected into the oil reservoir.However, in other applications the seawater is pretreated to avoidnegative impacts on the oil reservoir, such as acidifying the oil, e.g.by hydrogen sulfide (H₂S), or blocking pores or small passages in thereservoir, e.g. by sulfates. To achieve the desired seawater quality,the seawater is passed through a series of ever-finer filters providinga microfiltration of the seawater. In addition, biological orelectrochemical processes can be used to pretreat the seawater. Usuallythe final step of the filtration is a nanofiltration, in particular toremove the sulfates from the seawater. Nanofiltration is a membranefiltration process requiring to supply the water to the membrane unitwith a pressure of typically 25-50 bar (2.5-5.0 MPa). Particularly forreverse osmosis filtration the required pressure can even be higher.After the nanofiltration process the seawater is supplied to the waterinjection pump, pressurized and injected into the subterranean region,where the oil reservoir is located. Thus, pretreating and injecting theseawater into the oil reservoir usually requires two pumps, namely amembrane feed pump for supplying the membrane filtration unit with theseawater and a water injection pump for suppling the filtered seawaterto the well for introducing the seawater into the oil reservoir.

SUMMARY

In view of an efficient exploitation of oil and gas fields, it has beenfound that there is an increasing demand for pumps and in particularwater injection pumps that can be installed directly on the sea groundin particular down to a depth of 100 m, down to 500 m or even down tomore than 1,000 m beneath the water surface. The design of such pumpscan be challenging, in particular because these pumps operate in adifficult subsea environment for a long time period with as littlemaintenance and service work as possible. This requires specificmeasures to minimize the amount of equipment involved and to optimizethe reliability of the pump. In view of water injection pumps deployedon the sea ground and the pretreatment of the seawater, the membranefeed pump might be dispensed with, if the seawater injection system isinstalled at such a depth that the ambient water pressure is sufficientto feed the membrane filtration unit. For example, at 500 m below thewater surface the hydrostatic pressure of the seawater is already about50 bar, which might be high enough to feed the membrane filtration unit.

Water injection pumps for subsea applications have to deliver quite ahigh pressure increase, which might be 200 bar or even more. Forcentrifugal pumps this requires high speed but relatively little speedvariation. For driving such pumps liquid filled, or flooded inductionmotors or permanent magnet motors. Due to the liquid filling, theviscous losses caused by the viscous drag are considerably high and theyincrease approximately with the cube of the motor speed. Indeed, theseviscous losses limit the motor speed to about 6000 rpm. For even higherspeeds the viscous losses become too high to allow for an economical andefficient operation of the pump. Therefore, it has been proposed, forexample in WO 2016/189397, to couple the motor and the pump unit by acombination of a hydraulic coupling in series with a magnetic coupling.The hydraulic coupling is able to increase the operating speed of thepump as compared to the motor speed. A liquid filled electrical motor isconnected to a hydraulic coupling and a magnetic coupling driver sectionwhich are arranged in a hermetically sealed container. The container isfilled with a cooling and lubricating fluid which is circulated throughthe container and an externally arranged cooling coil. A magneticcoupling follower driven by the magnetic coupling driver drives the pumpimpeller(s). By this design the system hermetically separates the pumpedprocess fluid from the cooling and lubricating fluid. The systemrequires a specific cooling and lubricating fluid as well as themagnetic coupling to avoid a penetration of the process fluid into thedriving unit.

It has been found that for subsea installations on the sea ground thereliability of a pump and the minimization of wear and degradationwithin the pump are of utmost importance.

In WO 2018/077527 a centrifugal pump is disclosed, which is a multiphaseprocess fluid pump, and which is suited as a subsea pump forinstallation on the sea ground. The pump comprises a pump unit with apump shaft and impellers as well as a drive unit with a drive shaft fordriving the rotation of the pump shaft and the impellers. The driveshaft is directly coupled to the pump shaft by means of a hydrodynamictorque converter, i.e. without an additional magnetic coupling.

It is an object of the invention to propose an improved centrifugal pumpwith a pump unit and a drive unit, wherein the drive unit is coupled tothe pump unit by means of a hydrodynamic coupling. The pump shall besuited for being configured for subsea applications and for deploymenton the sea ground.

The subject matter of the invention satisfying these objects ischaracterized by the features described herein.

Thus, according to the invention, a centrifugal pump for conveying aprocess fluid is proposed, having a pump unit, a drive unit for drivingthe pump unit, a pump inlet for receiving the process fluid, and a pumpoutlet for discharging the process fluid, wherein the pump unitcomprises at least one impeller for conveying the process fluid from thepump inlet to the pump outlet, and a pump shaft, on which each impelleris mounted, wherein the drive unit comprises a drive shaft for drivingthe pump shaft, and an electric motor for rotating the drive shaft aboutan axial direction, wherein a plurality of bearings is provided forsupporting the pump shaft and the drive shaft, wherein a hydrodynamiccoupling having a casing is provided for hydrodynamically coupling thedrive shaft to the pump shaft by a transmission fluid, and wherein atleast one of the plurality of bearings is arranged in the casing of thehydrodynamic coupling.

Thus, at least one bearing of the pump shaft or the drive shaft isintegrated into the hydrodynamic coupling. By integrating at least oneof the bearings for supporting the drive shaft and the pump shaft intothe casing of the hydrodynamic coupling the pump is considerably morecompact because the distance in axial direction between the pump unitand the drive unit is reduced.

Preferably the hydrodynamic coupling comprises a pump wheel connected tothe drive shaft, a turbine wheel connected to the pump shaft, and astator arranged between the pump wheel and the turbine wheel for guidingthe transmission fluid. Thereby the hydrodynamic coupling can be ahydrodynamic torque converter. This enables the pump shaft to be capableof rotating with a higher speed (as measured in rounds per minute (rpm))than the drive shaft. Thus, it is either possible to drive the pumpshaft with a higher speed, e.g. with 7800 rpm instead of 6000 rpm or toreduce the speed of the drive shaft. Both measures increase theefficiency of the pump. In addition, since the torque convertor can beused for the speed control of the pump or for a self-regulation of therotational speed of the pump impeller, it is also possible to replace avariable frequency drive (VFD) with a less expensive drive, e.g. asingle speed electric drive, or to use a VFD that has a narrower rangefor the variation of the frequency of the drive shaft rotation. Bothmeasures constitute a considerable advantage from the economicalperspective.

According to a preferred embodiment the plurality of bearings comprisesa first radial pump bearing for supporting the pump shaft, wherein thefirst radial pump bearing is arranged between the pump unit and a driveend of the pump shaft, and wherein the first radial pump bearing isarranged in the casing of the hydrodynamic coupling. By integrating thefirst radial pump bearing in the hydrodynamic coupling the overhang ofthe pump shaft is considerably reduced, because the center of gravity ofthe coupling between the pump shaft and the drive shaft is much closerto the first radial pump bearing. This measure increases the stabilityof the pump shaft during rotation and reduces susceptibility tounbalance induced by rotating overhung masses.

For the same reason it is preferred that the plurality of bearingscomprises a first radial drive bearing for supporting the drive shaft,wherein the first radial drive bearing is arranged between the pump unitand the electric motor, and wherein the first radial drive bearing isarranged in the casing of the hydrodynamic coupling. By integrating thefirst radial drive bearing in the hydrodynamic coupling the overhang ofthe drive shaft is considerably reduced, because the center of gravityof the coupling between the drive shaft and the pump shaft is muchcloser to the first radial drive bearing. This measure increases thestability of the drive shaft during rotation and reduces susceptibilityto unbalance induced by rotating overhung mass.

It is more preferred that both the first radial pump bearing, and thefirst radial drive bearing are arranged in the casing of thehydrodynamic coupling. Beside the reduced overhang of both the driveshaft and the pump shaft, this design is very compact and reduces theoverall dimension of the pump.

According to a preferred embodiment, the centrifugal pump is a processfluid lubricated pump, and has a common housing, wherein the pump unitand the drive unit are arranged in the common housing, and wherein theplurality of bearings is configured to receive the process fluid aslubricant and coolant. By using the process fluid as coolant and aslubricant for the plurality of bearing further reduces the complexity ofthe pump.

Furthermore, it is preferred, that the hydrodynamic coupling isconfigured to receive the process fluid as the transmission fluid. Byusing the process fluid as the transmission fluid for hydrodynamicallycoupling the drive shaft to the pump shaft, it is no longer necessary toarrange the hydrodynamic coupling within a container that ishermetically sealed with respect to the process fluid. There is no needfor a specific transmission fluid. It is only the process fluid which isused for hydrodynamically coupling the drive unit to the pump unit.Furthermore, there is no need for an additional magnetic coupling. Thereis only one coupling between the electric motor and the pump unit, whichconsiderably reduces the transmission losses. Thereby the complexity ofthe pump is considerably reduced, and the reliability is increased.

According to a particularly preferred embodiment, the pump is aseal-less pump without a mechanical seal. A mechanical seal is usuallyused for the sealing the rotating shaft of a pump and prevents theleakage of the process fluid along the shaft of the pump. Typically, amechanical seal comprises a stator and a rotor. The rotor is connectedin a torque-proof manner with the shaft of the pump and the stator isfixed with respect to the pump housing such that the stator is securedagainst rotation. During rotation of the shaft the rotor is in slidingcontact with the stator thus performing the sealing action. Althoughsuch mechanical seals are widely spread within the technology ofcentrifugal pumps, they are somewhat problematic for subsea applicationsbecause they are quite complicated and usually require additionalequipment, which is often considered as a drawback for subseaapplications. Therefore, it is preferred that the pump according to theinvention is designed as a seal-less pump, i.e. a pump that has nomechanical seal. In many applications this requires that the pump unitand the drive unit are flooded with the process fluid. The advantage ofthe seal-less pump is the simpler design of the pump. In addition, theprocess fluid itself can be used for cooling and lubricating componentsof the pump, e.g. the bearing units of the pump shaft and the drive unitof the pump. Preferably the seal-less pump is configured and operatedwithout a barrier fluid system, i.e. during operation there is nobarrier fluid required or provided.

Preferably, the pump comprises a balance line configured for therecirculation of process fluid from a high pressure side to a lowpressure side of the pump. By providing the balance line forrecirculating an amount of the process fluid from the high pressure sideto the low pressure side, it is possible to guide the process fluid toall bearings for cooling and lubricating the bearings.

Furthermore, it is preferred that the electric motor is configured to bepassed through and cooled by the process fluid, so that the electricmotor is also cooled by the process fluid.

According to a preferred design the balance line is arranged andconfigured to receive process fluid discharged from the drive unit.Thus, the process fluid, which is recirculated from the high pressureside, passes through all bearings that are arranged at the high pressureside, optionally through the hydrodynamic coupling, through the electricmotor and then enters the balance line for being recirculated to the lowpressure side. This design renders possible that only the pressuregenerated by the pump unit is used to circulate the process fluidthrough the pump for cooling and lubricating in particular the pluralityof bearings as well as the drive unit.

According to another preferred design the pump has an external coolingloop for cooling and lubricating the drive unit and the plurality ofbearings, the external cooling loop comprising a heat exchanger forcooling the process fluid, wherein the heat exchanger is arrangedoutside the common housing and configured to receive process fluid fromthe common housing and to recirculate the process fluid to the commonhousing.

For moving the process fluid through the external cooling loop, acirculation impeller or a plurality of circulation impellers can beprovided. The circulation impeller for the external cooling circuit ispreferably rotated by the drive unit and can be arranged on top of thedrive unit. The drive unit drives the circulation impeller, whichcirculates the process fluid through the heat exchanger and the bearingunits. The heat exchanger can be a coil surrounding the common housingof the pump.

According to still another preferred design the pump unit comprises anintermediate take-off connected to a cooling loop, wherein theintermediate take-off is configured to supply the process fluid to thecooling loop with a pressure that is larger than the pressure of theprocess fluid at the pump inlet, and wherein the cooling loop isconfigured to supply process fluid to at least one bearing of theplurality of bearings and/or to the drive unit. Thus, the pressure forcirculating the process fluid e.g. through plurality of bearings istaken from the pump unit itself by means of the intermediate take-off.

It is also possible that a part of the process fluid, which is used astransmission fluid in the hydrodynamic coupling, is extracted e.g. fromthe casing of the hydrodynamic coupling and supplied to the coolingloop.

Preferably, the pump comprises at least one balance drum or center bushor throttle bush fixedly connected to the pump shaft and defining afront side facing the pump unit and a back side, and further comprises arelief passage, which is provided between the balance drum and astationary part configured to be stationary with respect to the commonhousing, wherein the relief passage is extending in axial directionalong the balance drum from the front side to the back side. The atleast one balance drum or center bush or throttle bush at leastpartially balances the axial thrust that is generated by the impeller(s)during operation of the pump.

According to a preferred design the balance drum is arranged between thepump unit and the hydrodynamic coupling and preferably at the highpressure side so that the front side is exposed to the high pressure ora pressure that approximately equals the high pressure. Since the frontside is exposed essentially to the high pressure, a pressure drop existsover the balance drum resulting in a force which counteracts the axialthrust generated by the impeller(s) during operation of the pump.

It is also possible to provide the balance drum at the low pressure sideso that the front side is exposed to the low pressure prevailing at thepump inlet.

Of course, it is also possible to provide more than one balance drumand/or other devices for balancing the axial thrust generated by theimpeller(s), such as a center bush or a throttle bush (also referred toas throttle sleeve). In particular, it is possible to provide twobalance drums on the pump shaft, each delimiting a respective reliefpassage between the respective balance drum and a stationary part.Preferably the two balance drums are arranged on both sides of the pumpunit, i.e. at the high pressure side and at the low pressure side, suchthat the pump unit is interposed between the two balance drums.Providing two or even more balance drums or other balancing devices canbe advantageous with respect to the rotor dynamic, which can beconsiderably improved. The rotor comprises all the rotating parts of thepump unit, namely the pump shaft, all impellers and the balance drum(s)fixed to the pump shaft. In particular, the improved rotor dynamicresults from an increased rotor stability. Each balance drum contributesto the rotor stability and enhances the rotor stability. An increasedrotor stability can result in a considerably reduced risk of wear, inparticular in the bearing units supporting the pump shaft. In addition,the improved rotor dynamic also enhances the reliability and reduces thesusceptibility to failure.

According to a preferred design the pump is a vertical pump with thepump shaft extending in the direction of gravity, and wherein the driveunit is arranged on top of the pump unit.

According to a preferred application the pump is configured forinstallation on a sea ground. The pump can be installed at a depth ofdown to 100 m, down to 500 m or even down to more than 1,000 m beneaththe water surface.

According to a preferred embodiment the pump is a water injection pumpfor injecting seawater into a subterranean region.

Further advantageous measures and embodiments of the invention willbecome apparent from the dependent claims.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be explained in more detail hereinafter withreference to the drawings.

FIG. 1 is a schematic cross-sectional view of a first embodiment of acentrifugal pump according to the invention,

FIG. 2 is a schematic representation of an embodiment of the drive unitwith the drive bearings and a hydrodynamic coupling,

FIG. 3 is a cross-sectional view of an embodiment of the hydrodynamiccoupling,

FIG. 4 is a schematic cross-sectional view of the first embodiment withanother embodiment of a cooling loop, and

FIG. 5 is a schematic cross-sectional view of a second embodiment ofcentrifugal pump according to the invention.

DETAILED DESCRIPTION OF THE EMBODIMENTS

FIG. 1 shows a schematic cross-sectional view of a first embodiment of acentrifugal pump according to the invention, which is designated in itsentity with reference numeral 1. By way of example, the pump 1 isdesigned as a process fluid lubricated pump for conveying a processfluid and has a common housing 2, a pump unit 3 and a drive unit 4. Boththe pump unit 3 and the drive unit 4 are arranged within the commonhousing 2. The common housing 2 is designed as a pressure housing, whichis able to withstand the pressure generated by the pump 1 as well as thepressure exerted on the pump 1 by the environment. The common housing 2can comprise several housing parts, e.g. a pump housing and a drivehousing, which are connected to each other to form the common housing 2surrounding the pump unit 3 and the drive unit 4. The common housing 2is a hermetically sealed pressure housing preventing any leakage to theexternal environment.

In the following description reference is made by way of example to theimportant application that the centrifugal pump is configured as aprocess fluid lubricated pump 1, which is designed and adapted for beingused as a subsea water injection pump 1 in the oil and gas industry, inparticular for injecting water into a subterranean oil and/or gasreservoir to increase recovery of hydrocarbons from the subterraneanregion. By injecting the water into the reservoir, the hydrocarbons areforced to flow towards and out of the production well. Accordingly, theprocess fluid that is conveyed by the pump 1 is water and especiallyseawater. The process fluid lubricated pump 1 is in particularconfigured for installation on the sea ground, i.e. for use beneath thewater surface, in particular down to a depth of 100 m, down to 500 m oreven down to more than 1000 m beneath the water surface of the sea.

It goes without saying that the invention is not restricted to thisspecific example but is related to centrifugal pumps in general. Theinvention can be used for many different applications, e.g. for suchapplications where the pump 1 is installed at locations, which aredifficult to access. By way of example, the pump 1 according to theinvention is designed as a water injection pump. Even if preferred, thepump 1 is not necessarily configured for deployment on the sea ground orfor subsea applications, but can also be configured for top sideapplications, e.g. for an installation on shore or on an oil platform,in particular on an unmanned platform. In addition, the pump 1 accordingto the invention can also be used for applications outside the oil andgas industry.

The term “process fluid lubricated pump” refers to pumps, where theprocess fluid that is conveyed by the pump 1 is used for the lubricationand the cooling of components of the pump, e.g. bearing units. A processfluid lubricated pump 1 does not require a specific barrier fluiddifferent from the process fluid to avoid leakage of the process fluide.g. into the drive unit 4. In addition, a process fluid lubricated pump1 does not require a lubricant different from the process fluid for thelubrication of the pump components. In the following descriptionreference is made by way of example to the important application thatthe process fluid is water, in particular seawater. The term seawatercomprises raw seawater, purified seawater, pretreated seawater, filteredseawater, in particular microfiltered seawater and nanofilteredseawater. Of course, the pump 1 according to the invention can also beconfigured for conveying other process fluids than water or seawater.

It goes without saying that the invention is not restricted to processfluid lubricated pumps but relates to centrifugal pumps in general.

The common housing 2 of the pump 1 comprises a pump inlet 21, throughwhich the process fluid enters the pump 1, and a pump outlet 22 fordischarging the process fluid with an increased pressure as compared tothe pressure of the process fluid at the pump inlet 21.

Typically, the pump outlet 22 is connected to a pipe (not shown) fordelivering the pressurized process fluid to a well, in which the processfluid is injected. The pressure of the process fluid at the pump outlet22 is referred to as ‘high pressure’ whereas the pressure of the processfluid at the pump inlet 21 is referred to as ‘low pressure’. A typicalvalue for the difference between the high pressure and the low pressureis for example 100 to 200 bar (10-20 MPa).

The pump unit 3 further comprises a pump shaft 5 extending from a driveend 51 to a non-drive end 52 of the pump shaft 5. The pump shaft 5 isconfigured for rotating about an axial direction A, which is defined bythe longitudinal axis of the pump shaft 5.

The pump unit 3 further comprises at least on impeller 31 fixedlymounted on the pump shaft 5 and configured for increasing the pressureof the pressure fluid from the low pressure to the high pressure.Preferably the pump unit 3 comprises a plurality of impellers 31 mountedin series on the pump shaft 5 in a torque proof manner. FIG. 1 shows anexample where the pump unit 3 comprises ten impellers 31 arranged inseries on the pump shaft 5.

In other embodiments the pump unit can comprise a first set of impellersand a second set of impellers which are arranged in a back-to-backarrangement.

The drive unit 4, which will be explained in more detail hereinafter, isconfigured to exert a torque on the drive end 51 of the pump shaft 5 fordriving the rotation of the pump shaft 5 and the impellers 31 about theaxial direction A.

The process fluid lubricated pump 1 is configured as a vertical pump 1,meaning that during operation the pump shaft 5 is extending in thevertical direction, which is the direction of gravity. Thus, the axialdirection A coincides with the vertical direction.

A direction perpendicular to the axial direction is referred to asradial direction. The term ‘axial’ or ‘axially’ is used with the commonmeaning ‘in axial direction’ or ‘with respect to the axial direction’.In an analogous manner the term ‘radial’ or ‘radially’ is used with thecommon meaning ‘in radial direction’ or ‘with respect to the radialdirection’. Hereinafter relative terms regarding the location like“above” or “below” or “upper” or “lower” or “top” or “bottom” refer tothe usual operating position of the pump 1. FIG. 1, FIG. 4 and FIG. 5show different embodiments and variants of the pump 1 in theirrespective usual operating position.

Referring to this usual orientation during operation and as shown inFIG. 1 the drive unit 4 is located above the pump unit 3. However, inother embodiments the pump unit 3 can be located on top of the driveunit 4. In other embodiments the pump 1 can be configured as ahorizontal pump, i.e. for a horizontal arrangement with the pump shaft 5extending horizontally during operation of the pump.

The pump inlet 21 is arranged at the lower end of the pump unit 3, andthe pump outlet 22 is located at the upper end of the pump unit 3.

In other embodiments the pump inlet 21 can be arranged at the upper endof the pump unit 3, and the pump outlet 22 can be located at the lowerend of the pump 1.

The pump 1 comprises a plurality of bearings. A first radial pumpbearing 53, a second radial pump bearing 54 and an axial pump bearing 55are provided for supporting the pump shaft 5. The first radial pumpbearing 53, which is the upper one, is arranged adjacent to the driveend 51 of the pump shaft 5 between the pump unit 3 and the drive unit 4.The second radial pump bearing 54, which is the lower one, is arrangedbetween the pump unit 3 and the non-drive end 52 of the pump shaft 5 orat the non-drive end 52. The axial pump bearing 55 is arranged betweenthe pump unit 3 and the first radial pump bearing 53. The pump bearings53, 54, 55 are configured to support the pump shaft 5 both in axial andradial direction. The radial pump bearing 53 and 54 are supporting thepump shaft 5 with respect to the radial direction, and the axial bearing55 is supporting the pump shaft 5 with respect to the axial direction A.The first radial pump bearing 53 and the axial pump bearing 55 arearranged such that the first radial pump bearing 53 is closer to thedrive unit 4 and the axial pump bearing 55 is facing the pump unit 3. Ofcourse, it is also possible, to exchange the position of the firstradial pump bearing 53 and the axial pump bearing 55, i.e. to arrangethe first radial pump bearing 53 between the axial pump bearing 55 andthe pump unit 3, so that the axial pump bearing 55 is closer to thedrive unit 4.

A radial bearing, such as the first or the second radial pump bearing 53or 54 is also referred to as a “journal bearing” and an axial bearing,such as the axial pump bearing 55, is also referred to as an “thrustbearing”. The first radial pump bearing 53 and the axial pump bearing 55can be separate bearings, but it is also possible that the first radialpump bearing 53 and the axial pump bearing 55 are a single combinedradial and axial bearing supporting the pump shaft 5 both in radial andin axial direction.

The second radial pump bearing 54 is supporting the pump shaft 5 inradial direction. In the embodiment shown in FIG. 1, there is no axialpump bearing provided at the non-drive end 52 of the pump shaft 5. Ofcourse, in other embodiments it is also possible that an axial pumpbearing for the pump shaft 5 is provided at the non-drive end 52. Inembodiments, where an axial pump bearing is provided at the non-driveend 52, a second axial pump bearing can be provided at the drive end 51or the drive end 51 can be configured without an axial pump bearing.

Preferably the radial pump bearings 53 and 54 as well as the axial pumpbearing 55 are hydrodynamic bearings, and even more preferred as tiltingpad bearings 53, 54 and 55, respectively. Specifically preferred atleast the first radial pump bearing 53 and the second radial pumpbearing 54 are each configured as a radial tilting pad bearing. Ofcourse, it is also possible that the first radial pump bearing 53 andthe second radial pump bearing 54 are each fixed multilobe hydrodynamicbearings.

Preferably, the pump 1 comprises at least one balancing device for atleast partially balancing the axial thrust that is generated by theimpellers 31 during operation of the pump 1. The balancing device cancomprise a balance drum 7 or a center bush 7 or a throttle bush 7 (alsoreferred to as throttle sleeve). The first embodiment of the pump 1comprises a balance drum 7 for at least partially balancing the axialthrust that is generated by the impellers 31. The balance drum 7 isfixedly connected to the pump shaft 5 in a torque proof manner. Thebalance drum 7 is arranged above the upper end of the pump unit 3,namely between the pump unit 3 and the drive end 51 of the pump shaft 5,more precisely between the upper end of the pump unit 3 and the axialpump bearing 55. The balance drum 7 defines a front side 71 and a backside 72. The front side 71 is the side facing the pump unit 3 and theimpellers 31. The back side 72 is the side facing the axial pump bearing55 and the drive unit 4. The balance drum 7 is surrounded by astationary part 26, so that a relief passage 73 is formed between theradially outer surface of the balance drum 7 and the stationary part 26.The stationary part 26 is configured to be stationary with respect tothe common housing 2. The relief passage 73 forms an annular gap betweenthe outer surface of the balance drum 7 and the stationary part 26 andextends from the front side 71 to the back side 72. The front side 71 isin fluid communication with the pump outlet 22, so that the axialsurface of the balance drum 7 facing the front side 71 is exposedessentially to the high pressure prevailing at the pump outlet 22 duringoperation of the pump 1. Of course, due to smaller pressure lossescaused by the fluid communication between the pump outlet 22 and thebalance drum 7 the pressure prevailing at the axial surface of thebalance drum 1 facing the front side 71 can be somewhat smaller than thehigh pressure. However, the considerably larger pressure drop takesplace over the balance drum 7. At the back side 72 an intermediatepressure prevails during operation of the pump 1. The intermediatepressure has a value between the low pressure at the pump inlet 21 andthe high pressure at the pump outlet 22. In the embodiment shown in FIG.1 the intermediate pressure is somewhat larger than the low pressure dueto the pressure drop over the balance line 9.

Since the front side 71 is exposed essentially to the high pressure atthe pump outlet 22, a pressure drop exists over the balance drum 7resulting in a force that is directed upwardly in the axial direction Aand therewith counteracts the downwardly directed axial thrust generatedby the impellers 31 during operation of the pump 1.

In other embodiments (not shown) a further balance drum is arrangedbelow the lower end of the pump unit 3, namely between the pump unit 3and the non-drive end 52 of the pump shaft 5, more precisely between thelower end of the pump unit 3 and the second radial pump bearing 54. Instill other embodiments a balance drum is provided only at the lower endof the pump 1, between the pump unit 3 and the second radial pumpbearing 54 at the non-drive end 52 of the pump shaft 5 and no balancedrum is provided above the pump unit 3 near the drive end of the pumpshaft 5. It is also possible to arrange a balancing device between twoadjacent impellers 31, for example between a first set of impellers anda second set of impellers. The first set of impellers and the second setof impellers can be configured in a back-to-back arrangement.

In the embodiment shown in FIG. 1 a balance line 9 is provided forrecirculating the process fluid from the high pressure side to the lowpressure side. In particular, the balance line 9 connects the back side72 with the low pressure side of the pump 1, where the low pressure,i.e. the pressure at the pump inlet 21 prevails. Thus, a part of thepressurized process fluid passes from the high pressure side through therelief passage 73 to the back side 72, enters the balance line 9 and isrecirculated to the low pressure side of the pump 1. The balance line 9constitutes a flow connection between the back side 72 and the lowpressure side at the pump inlet 21. The balance line 9 can bearranged—as shown in FIG. 1—outside the common housing 2. In otherembodiments the balance line 9 can be designed as internal linecompletely extending within the common housing 2.

The balance line 9 extends from a first port 91 at the back side 72 to asecond port 92, which is in fluid communication with the low pressureside of the pump, e.g. with the pump inlet 21. As indicated in FIG. 1the second port 92 can be arranged between the second radial pumpbearing 54 and the first one of the impellers 31. The first and thesecond port 91, 92 are arranged at the common housing 2. Thus, duringoperation of the pump 1 the process fluid can flow from the back side 72through the balance line 9 to the low pressure side of the pump 1.Therefore, the pressure prevailing at the back side 72, namely theintermediate pressure is essentially the same—apart from a minorpressure drop caused by the balance line 9—as the low pressureprevailing at the pump inlet 21.

The process fluid lubricated pump 1 is designed as a seal-less pump. Aseal-less pump 1 is a pump that has no mechanical seals for the sealingof the rotating pump shaft 5. A mechanical seal is a seal for a rotatingshaft comprising a rotor fixed to the shaft and rotating with the shaftas well as a stationary stator fixed with respect to the housing. Duringoperation the rotor and the stator are sliding along each other—usuallywith a liquid there between—for providing a sealing action to preventthe process fluid from escaping to the environment or entering the driveof the pump. The seal-less pump 1 shown in FIG. 1 has no such mechanicalseals. The process fluid is deliberately allowed to enter the drive unit4 and is used for cooling and lubricating components of the pump 1 suchas the pump bearings 53, 54 and 55.

FIG. 2 shows a schematic representation of an embodiment of the driveunit 4 with drive bearings 43, 44 and 45 in more in detail. The driveunit 4 comprises an electric motor 41 and a drive shaft 42 extending inthe axial direction A. For supporting the drive shaft 42 a first radialdrive bearing 43, a second radial drive bearing 44 and an axial drivebearing 45 are provided, wherein the second radial drive bearing 44 andthe axial drive bearing 45 are arranged above the electric motor 41 withrespect to the axial direction A, and the first radial drive bearing 43is arranged below the electric motor 41. The electric motor 41, which isarranged between the first and the second radial drive bearing 43, 44,is configured for rotating the drive shaft 42 about the axial directionA. The drive shaft 42 is connected to the drive end 51 of the pump shaft5 by means of a hydrodynamic coupling 8 for transferring a torque to thepump shaft 5. Preferably the hydrodynamic coupling 8 is configured as atorque converter 8 for hydrodynamically coupling the drive shaft 42 tothe pump shaft 5 in such a manner that the pump shaft 5 can rotate at ahigher speed (rounds per minute, rpm) than the drive shaft 42.

The drive bearings 43, 44 and 45 are configured to support the driveshaft 42 both in radial direction and in the axial direction A. Thefirst and the second radial drive bearing 43, 44 support the drive shaft42 with respect to the radial direction, and the axial drive bearing 45supports the drive shaft 42 with respect to the axial direction A. Thesecond radial drive bearing 44 and the axial drive bearing 45 arearranged such that the second radial drive bearing 44 is arrangedbetween the axial bearing 45 and the electric motor 41.

Of course, it is also possible, to exchange the position of the secondradial drive bearing 44 and the axial drive bearing 45.

The second radial drive bearing 44 and the axial drive bearing 45 can beseparate bearings, but it is also possible that the second radial drivebearing 44 and the axial drive bearing 45 are a single combined radialand axial bearing supporting the drive shaft 42 both in radial and inaxial direction A.

The first radial drive bearing 43 is arranged below the electric motor41 and supports the drive shaft 42 in radial direction. In theembodiment shown in FIG. 2, there is no axial bearing arranged below theelectric motor 41. Of course, it is also possible that an axial drivebearing for the drive shaft 42 is—alternatively or additionally—arrangedbelow the electric motor 41, i.e. between the electric motor 41 and thehydrodynamic coupling 8.

The electric motor 41 of the drive unit 4 comprises an inwardly disposedrotor 412, which is connected to the drive shaft 42 in a torque proofmanner, as well as an outwardly disposed motor stator 411 surroundingthe rotor 412 with an annular gap 413 between the rotor 412 and themotor stator 411. The rotor 412 can constitute a part of the drive shaft42 or is a separate part, which is rotationally fixedly connected to thedrive shaft 42, so that the rotation of the rotor 412 drives the driveshaft 42. The electric motor 41 can be a cable wound motor. In a cablewound motor, the individual wires of the motor stator 411, which formthe coils for generating the electromagnetic field(s), are eachinsulated, so that the motor stator 411 can be flooded even with anelectrically conducting fluid, e.g. raw seawater. The cable wound motordoes not require a dielectric fluid for cooling the motor stator 411.Alternatively, the electric motor 41 can be a canned motor. When theelectric drive 41 is configured as a canned motor, the annular gap 413is radially outwardly delimited by a can (not shown) that seals themotor stator 411 hermetically with respect to the rotor 412 and the gap413. Thus, any process fluid flowing through the gap 413 cannot enterthe motor stator 411. When the electric motor 41 is designed as a cannedmotor a dielectric cooling fluid different from the process fluid, canbe circulated through the hermetically sealed motor stator 411 forcooling the motor stator 411.

Preferably, the electric motor 41 is a permanent magnet motor or as aninduction motor. To supply the electric motor 41 with energy, a powerpenetrator (not shown) is provided at the common housing 2 for receivinga power cable (not shown) that supplies the electric motor 41 withpower.

The electric motor 41 can be designed to operate with a variablefrequency drive (VFD), in which the speed of the motor, i.e. thefrequency of the rotation, is adjustable by varying the frequency and/orthe voltage supplied to the electric motor 41. However, it is alsopossible that the electric motor 41 is configured differently, forexample as a single speed or single frequency drive.

FIG. 3 shows a first embodiment of the hydrodynamic coupling 8 which isa torque converter 8, for hydrodynamically coupling the drive shaft 42to the pump shaft 5 in such a manner that the pump shaft 5 can rotatefaster than the drive shaft 42, e.g. by a factor of 1.3 Thus, when thedrive shaft 42 rotates for example at 6000 rpm the pump shaft 5 rotatesat 7800 rpm.

The torque converter 8 comprises a casing 81 for receiving atransmission fluid, a pump wheel 82 connected in a torque proof mannerto the drive shaft 42 of the drive unit 4, a turbine wheel 83 connectedin a torque proof manner to the pump shaft 5 of the pump unit 3, and astator 84 for guiding the transmission fluid from the turbine wheel 83back to the pump wheel 82 in a manner that is as such known in the art.

The stator 84 is connected to the casing 81, which is a liquid tightcasing 81 in which the transmission fluid is contained. During operationthe drive shaft 42 rotates the pump wheel 82 which acts on andaccelerates the transmission fluid thereby transforming mechanicalenergy in flow energy. The transmission fluid transfers the energy tothe turbine wheel 83 that transforms the flow energy back to mechanicalenergy and drives the pump shaft 5. The transmission fluid is thenrecirculated to the pump wheel 82 being guided and diverted by thestator 84.

In the embodiment shown in FIG. 3 the torque converter 8 comprises aplurality of fixed guide vanes 85, i.e. the position of the guide vanes85 cannot be changed, but the guide vanes 85 are fixed with respect tothe casing 81 and cannot be moved relative to the casing 81.

Of course, in other embodiments it is also possible to configure thetorque converter with guide vanes, which are movable with respect to thecasing 81, i.e. the guide vanes are configured as adjustable guide vaneswith which the streaming of the transmission fluid can be changed. Theposition of the adjustable guide vanes can be changed by means of anadjustment device. By pivoting the guide vanes, the incident flow of theguide vanes can be changed, whereby the flow of the transmission fluidin the casing is modified. Thereby the ratio of the torque transmittedto the pump shaft and the torque transmitted from the drive shaft iscontrollable.

Providing the torque convertor 8 for coupling the dive shaft 42 with thepump shaft 5 instead of e.g. a mechanical coupling of the drive shaft 42and the pump shaft 5 has the important advantage that the pump shaft 5can be rotated at a higher speed than the drive shaft. This increasesthe efficiency of the pump 1, because the impeller(s) 31 are rotatingfaster and/or because the viscous losses in the liquid filled electricmotor 41 are considerably reduced. Furthermore, by the hydrodynamiccoupling the drive shaft 42 and the pump shaft 5 are torsionallydecoupled, reducing e.g. impact from drive unit 4 torsional shock loadslike short circuit or startup.

It is a further advantage that the torque convertor 8 can also be usedfor the speed control of the pump 1 or for a self-regulation of thepumping process. This is in particular applicable when the torqueconvertor is designed with adjustable guide vanes, but also applicablewhen the torque convertor 8 has only fixed guide vanes 85 (as shown inFIG. 3) for the transmission fluid and no adjustable guide vanes. Usingthe torque converter 8 for the speed control of the pump shaft 5 renderspossible to operate the pump 1 without a variable frequency drive (VFD)but with a less complex and less expensive electric motor 41.Alternatively, it is possible to use a VFD with a considerably narrowerrange for the variation of the frequency. Both measures, configuring thepump 1 without a VFD or configuring the pump with a VFD having anarrower frequency range, reduce the costs and the complexity of thepump 1.

According to embodiments of the invention at least one of the pluralityof bearings 43, 44, 45, 53, 54, 55 is arranged in the casing 81 of thehydrodynamic coupling 8. In the embodiment shown in FIG. 3 and FIG. 2the first radial pump bearing 53 and the first radial drive bearing 43are both arranged in the casing 81 of the torque converter 8. Thus, thefirst radial pump bearing 53 and the first radial drive bearing 43 areintegrated in the torque converter 8, more precisely in the casing 81,which prevents the transmission fluid from escaping.

Configuring the torque converter 81 with the integrated first radialpump bearing 53 and with the integrated first radial drive bearing 43has the advantage that the overhang of both the pump shaft 5 and thedrive shaft 42 is considerably reduced, because the first radial pumpbearing 53 is located closer to the drive end 51 of the pump shaft 5 andthe first radial drive bearing 43 is located closer to the drive end ofthe drive shaft 42.

As can be best seen in FIG. 3, the first radial pump bearing 53 isarranged in the casing 81 and between the casing 81 and the turbinewheel 83, so that the first radial pump bearing 53 surrounds the turbinewheel 83. Of course, in another configuration the first radial pumpbearing 53 can also be arranged in the casing 81 such that the bearing53 surrounds the pump shaft 5. The first radial drive bearing 43 isarranged in the casing 81 and between the casing 81 and the drive shaft42, so that the first radial drive bearing 43 surrounds the drive shaft42. Of course, in another configuration the first radial drive bearing43 can also be arranged in the casing 81 such that the first radialdrive bearing 43 surrounds the pump wheel 82.

Although it is preferred that both the first radial pump bearing 53 andthe first radial drive bearing 43 are arranged in the casing 81, inother embodiments only the first radial drive bearing 43 or only thefirst radial pump bearing 53 is arranged in the casing 81, so that onlyone bearing 43 or 53 is integrated in the torque converter 8.

In still other embodiments it is also possible that at least one of theaxial pump bearing and the axial drive bearing is arranged in the casing81 of the torque converter 8. Exemplary, a few configurations are listedhere:

-   only the axial drive bearing 45 is arranged in the casing 81 of the    torque converter 8;-   only the axial drive bearing 45 and the first radial drive bearing    43 are arranged in the casing 81 of the torque converter 8;-   only the axial pump bearing 55 is arranged in the casing 81 of the    torque converter 8;-   only the axial pump bearing 55 and the first radial pump bearing 53    are arranged in the casing 81 of the torque converter 8;-   only the axial pump bearing 55, the first radial pump bearing 53 and    the first radial drive bearing 43 are arranged in the casing 81 of    the torque converter 8;-   the axial pump bearing 55, the axial drive bearing 45 the first    radial pump bearing 53 and the first radial drive bearing 43 are    arranged in the casing 81 of the torque converter 8.

Particularly preferred, the hydrodynamic coupling is configured toreceive the process fluid as the transmission fluid. Thus, the processfluid, e.g. water or seawater, is not only used for cooling andlubricating the bearings 43, 44, 45, 53, 54, 55 and the electric motor41 but also as the transmission fluid for the hydrodynamic coupling 8.

The casing 81 of the torque converter 8 comprises an intake 86 forsupplying the process fluid to the casing 81 of the torque convertor 8and a discharge 87 for the drain of the process fluid. As an alternativeor as a supplement the torque converter 8 can be configured such thatthe process fluid can enter and leave the casing 81, for example as aleakage flow, along the pump shaft 5 and/or the drive shaft 42. Duringoperation of the pump 1 the entire common housing 2 of the pump 1 isflooded with the process fluid, so that the process fluid can enter andleave the casing 81 of the torque converter 8 through the intake 86 andthe discharge 87 and/or as a leakage along the pump shaft 5 and thedrive shaft 42.

Since the torque converter 8 is immersed in the process fluid duringoperation of the pump 1, the process fluid can leak between the firstradial pump bearing 53 and the turbine wheel 83 or the pump shaft 5,respectively, into and out of the torque converter 8. Furthermore, theprocess fluid can leak between the first radial drive bearing 43 and thedrive shaft 42 or the pump wheel 82, respectively, into and out of thetorque converter 8.

In other embodiments the casing 81 has no intake 86 and no discharge 87,so that the process fluid can enter and leave the casing 81 only as aleakage along the pump shaft 5 and/or the drive shaft 42.

Depending on the pressure in the common housing 2 to which the exteriorof the casing 81 of the torque converter 8 is exposed, the torqueconvertor 8 can generate an additional driving force for recirculatingthe process fluid from the high pressure side to the low pressure sideof the pump 1 through the balance line 9.

During operation, the pump 1 is cooled and lubricated by the processfluid, e.g. seawater. In the first embodiment, shown in FIG. 1, anexternal cooling loop 10 enhances the cooling of the pump 1. Theexternal cooling loop 10 is also operated with the process fluid, e.g.seawater, as heat carrier. According to this embodiment, the externalcooling loop 10 comprises at least one circulation impeller 11 forcirculating the process fluid through the external cooling loop 10. Thecirculation impeller 11 is a different feature than the impellers 31 ofthe pump unit 3.

Since the process fluid constitutes the heat carrier, the externalcooling loop 10 can be designed as an open circuit, which receivesprocess fluid from the pump unit 3, and which delivers the process fluidto different locations of the pump 1. The circulation impeller 11 isdriven by the electric motor 41 and preferably by the drive shaft 42. Asshown in FIG. 1 the circulation impeller 11 can be arranged for exampleon top the electric motor 41, but other location are also possible. Forexample, the circulation impeller(s) 11 can also be arranged at one orat more of the following locations: the non-drive end of the drive shaft42, the drive end of the drive shaft 42, the drive end 51 of the pumpshaft 5, above the balance drum 7, above the first port 91 to thebalance line 9, at the non-drive end 52 of the pump shaft 5, below thesecond radial pump bearing 54.

The external cooling loop 10 further comprises a heat exchanger 12 forcooling the process fluid in the external cooling loop 10. The heatexchanger 12 is located outside the common casing 2. Preferably, theheat exchanger 12 is designed as a coil or a spiral that surrounds thecommon casing 2. In a subsea application, the seawater around the pump 1extracts heat from the coil-shaped heat exchanger 12 at the outside ofthe common housing 2 and therewith cools the process liquid in theexternal cooling loop 10. The flow of the process fluid in the externalcooling loop 10 is indicated in FIG. 1 with the dashed arrows. The heatexchanger 12 is in fluid communication with an exit 13 for receivingprocess fluid from the drive unit 4 as indicated by arrow C1. Moreprecisely the exit 13 is provided at the common housing 2 at a locationabove the drive unit 4 and above the axial drive bearing 45, so that theheat exchanger 12 receives process fluid that has passed through thedrive unit 4 and the drive bearings 43, 44, 45 and therewith cooled thedrive unit 4 and the drive bearings 43, 44, 45. In the heat exchanger 12the environment extracts heat from the process fluid and cools theprocess fluid. After having passed through the heat exchanger 12 thecooled process fluid is provided to several locations of the pump forcooling and lubricating the components. For each location a respectiveentrance 14, 15, 16 for the process fluid is disposed at the commonhousing 2. Downstream of the heat exchanger 12 a first part of thecooled process fluid, as indicated by arrow C2, is introduced throughentrance 14 directly into the drive unit 4 for cooling and lubricatingthe drive bearings 44, 45 as well as for cooling the electric motor 41.A second part of the cooled process fluid, as indicated by arrow C3, isintroduced through entrance 15 directly into the axial pump bearing unit55 for cooling and lubricating the axial pump bearing 55 and optionallyfor passing through the torque converter 8 as well as through the firstradial pump bearing 53 and the first radial drive bearing 43, which areboth arranged within the torque converter 8. A third part of the cooledprocess fluid, as indicated by arrow C4, is introduced through entrance16 directly into the second radial pump bearing 54 for cooling andlubricating the second radial pump bearing 54. The process fluid thatpasses through the electric motor 41 for cooling the electric motor isdirected through the annular gap 413 as indicated by the dashed arrowsC5 in FIG. 2. In case the motor stator 411 is flooded with the processfluid for cooling, e.g. when the electric motor 41 is a cable woundmotor, the process fluid is also directed through the motor stator 411as indicated by the dashed arrows C6 in FIG. 2.

FIG. 4 shows a different design for a cooling loop 10′ in across-sectional view similar to FIG. 1. This design does not require thecirculation impeller 11 but can also comprise a circulation impeller. Inthe configuration shown in FIG. 4 no circulation impeller is providedfor. According to this design of the cooling loop 10′, the pump unit 3comprises an intermediate take-off 310 connected to the cooling loop 10′for supplying the process fluid to the cooling loop 10′ as indicated bythe dashed arrow C7 in FIG. 4. The intermediate take-off 310 isconfigured to supply the process fluid to the cooling loop 10′ at apressure which is larger than the low pressure at the low pressure inlet21.

The cooling loop 10′ comprises a first branch 101 enabling a fluidcommunication between the intermediate take-off 310 and an entrance 17,through which the process fluid can enter the axial drive bearing 45 forcooling and lubricating as indicated by the dashed arrows C71 in FIG. 4.The process fluid that has passed through the axial drive bearing 45 isguided through the second radial drive bearing 44, the drive unit 4, thetorque converter 8 containing the first radial drive bearing 43 and thefirst radial pump bearing 53, and then through the axial pump bearing 55for cooling and lubricating these components as indicated by the dashedarrows C73 in FIG. 4. The process fluid that passed through the axialpump bearing 55 merges with the process fluid that passed along thebalance drum 7 and enters the balance line 9.

Optionally the first branch 101 of the cooling loop 10′ can comprise afirst flow restrictor 103, e.g. a throttle or an orifice, provided inthe first branch 101 to regulate the flow of process fluid that flowsthrough the entrance 17.

The cooling loop 10′ further comprises a second branch 102 enabling afluid communication between the intermediate take-off 310 and anentrance 18, through which the process fluid can enter the second radialpump bearing 54 for cooling and lubricating the second radial pumpbearing 54 as indicated by the dashed arrows C72 in FIG. 4. When theprocess fluid has passed through the second radial pump bearing 54 itmerges with the process fluid at the low pressure side at the pump inlet21.

Optionally the second branch 102 can comprise a second flow restrictor104, e.g. a throttle or an orifice, disposed in the second branch 102 toregulate the flow of process fluid that passes through the second radialpump bearing 54.

The intermediate take-off 310 can be arranged to receive the processfluid from one of the impellers 31. Thus, according to the design shownin FIG. 4 the driving force for circulating the process fluid throughthe cooling loop 10′ is generated by one or more of the impellers 31 ofthe pump unit 3. Preferably, the intermediate take-off 310 is configuredsuch, that the pressure of the process fluid in the first and the secondbranch 101 and 102 is at least as large as the pressure of the processfluid in the balance line 9. Even more preferred, the pressure of theprocess fluid in the first and the second branch 101 and 102 of thecooling loop 10′ is at least a few bar higher, for example 10-30 barhigher than the pressure in the balance line 9.

The first and the second branch 101 and 102 of the cooling loop can beinternal lines completely extending within the common casing 2. It isalso possible—as shown in FIG. 4—that the first and the second branch101 and 102 are external lines arranged outside the common housing 2. Ithas to be noted that the cooling loop 10′ can also comprise a heatexchanger in an analogous manner as explained for the heat exchanger 12shown in FIG. 1.

The operation of the first embodiment of the pump 1 according to theinvention will now be described referring to FIG. 1 to FIG. 3. Theprocess fluid entering the pump 1 through the pump inlet 21 with the lowpressure is pressurized by the action of the rotating impellers 31, andleaves the pump 1 through the pump outlet 22 with the high pressure asindicated in FIG. 1 and FIG. 4 by the large solid line arrows withoutreference numeral. The front side 71 below the balance drum 7 is influid communication with the pump outlet 22. Therefore, a part of thepressurized process fluid passes through the relief passage 73 to theback side 72 as indicated by arrows B1 in FIG. 1. At the back side 72the intermediate pressure prevails which is smaller than the highpressure due to the pressure drop over the balance drum 7. Thus, a forceis generated acting upon the pump shaft 5. The force is directedupwardly in axial direction A and therewith partially balancing theaxial thrust that is generated by the impellers 31 and that is directeddownwardly in axial direction A. At the back side 72 a part of theprocess fluid enters the balance line 9 through the first port 91, andanother part enters the pump axial bearing 55 and merges with theprocess fluid of the external cooling loop 10, which enters the commonhousing 2 through the entrance 15.

The process fluid flowing through the balance line 9 is recirculated tothe low pressure side of the pump 1 and merges with the process fluidthat has been introduced from the external cooling loop 10 throughentrance 16 into the second radial pump bearing 54.

The balance line 9 causes a small pressure drop so that the intermediatepressure at the back side 72 is somewhat larger than the low pressure atthe pump inlet 21.

The balance drum 7 at least partially compensates the axial thrust onthe pump shaft 5 that is generated by the rotating impellers 31. Even ifthe balance drum 7 does not completely balance the axial thrust, theload that has to be carried by the axial pump bearing 55 is considerablyreduced.

In other embodiments a further balance drum is disposed at the non-driveend 52 of the pump shaft 5. This might additionally increase thestability of the entire rotor device comprising the pump shaft 5, theimpellers 31 and the two balance drums. [0109] By the two balance drumsa rotordynamic vibrations of the lower part of the pump shaft 5, i.e.the part of the pump shaft 5 adjacent to the non-drive end 52 is evenmore reliably prevented or at least additionally reduced.

Only by way of example and for the better understanding the followingdifferent pressures can prevail at and in the pump 1: When, as anexample, the pump 1 is deployed at the sea ground in a depth of 250 mbelow the water surface, the low pressure prevailing at the pump inlet21 is e.g. 25 bar. The pump 1 can increase the pressure by 175 bar.Thus, the high pressure at the high pressure outlet 22 is 200 bar. Whenneglecting other minor pressure losses such as the pressure losses inthe balance line 9, the pressure drop over the balance drum 7 is roughly175 bar. Accordingly, the intermediated pressure prevailing at the backside 72 is approximately the low pressure, i.e. 25 bar.

The cooling and the lubricating of the pump 1 by the process fluid isachieved both by the flow through the balance line 9, which is driven bythe action of the impellers 31 and indicated by the arrows in solidlines in FIG. 1, and by the flow through the external cooling loop 10indicated by the arrows in dashed lines. Both the flows contribute tocool and lubricate the pump bearings 53, 54 and 55, the drive bearings43, 44 and 45 as well as the electric motor 41 with the process fluid.

FIG. 5 shows a schematic cross-sectional view of a second embodiment ofa centrifugal pump 1 according to the invention. The centrifugal pump 1is configured as a process fluid lubricated pump 1.

In the following description of the second embodiment of the centrifugalpump 1 only the differences to the first embodiment are explained inmore detail. The explanations with respect to the first embodiment arealso valid in the same way or in analogously the same way for the secondembodiment. Same reference numerals designate the same features thathave been explained with reference to the first embodiment orfunctionally equivalent features. In particular, the drive unitexplained with reference to FIG. 2 as well as the torque converterexplained with reference to FIG. 3 can also be used for the secondembodiment.

Compared to the first embodiment, it is the main difference, that thesecond embodiment of the pump 1 does not comprise an external coolingloop 10. The pump bearings 53, 54, 55 as well as the drive unit 4comprising the electric motor 41 as well as the drive bearings 43, 44,45 as well as the torque converter 8 are only cooled and lubricated bythe flow of process fluid, which is driven by the action of theimpellers 31 of the pump unit 3. However, as already stated, dependingon the respective local pressure in the common housing 2, the torqueconverter 8 can generate an additional driving force for recirculatingthe process fluid from the high pressure side to the low pressure sideof the pump 1 through the balance line 9.

The first port 91, to which the balance line 9 is connected forreceiving the process fluid, is arranged above the axial drive bearing45. The process fluid passing along the balance drum 7 through therelief passage 73 flows through the axial pump bearing 55, through thetorque converter 8 containing the first radial pump bearing 53 and thefirst radial drive bearing 43, through the drive unit 4, through thesecond radial drive bearing 44, and through the axial drive bearing 45.Above the axial drive bearing 45 the first port 91 is located formingthe entrance to the balance line 9 as indicated by the arrow B2 in FIG.5. Thus, the balance line 9 receives the process fluid that isdischarged from the drive unit 4 and has passed through the secondradial drive bearing 44 and the axial drive bearing 45. Channeling theprocess fluid through the bearings 55, 53, 43, 44, 45, the torqueconverter 8 and the drive unit 4 results in a pressure drop between theback side 72 and the port 91. The pressure drop can be a few bar, e.g.about 10 bar. Thus, at the first port 91 prevails a pressure, which issomewhat smaller than the intermediate pressure prevailing at thebackside 72 between the balance drum 7 and the axial pump bearing 55.

The second port 92, to which the balance line 9 is connected, isarranged below the second radial pump bearing 54 at the non-drive end 52of the pump shaft 5. Thus, the process fluid exiting the balance line 9and passing through the second port 92 is guided to pass through thesecond radial pump bearing 54 before the recirculated process fluidreaches the low pressure side adjacent to the pump inlet 21. Since theprocess fluid is directed from the second port 92 through the secondradial pump bearing 54, the pressure prevailing at the second port 92 issomewhat larger than the low pressure, because there is an additionalpressure drop over the second radial pump bearing 54 for example apressure drop of approximately four bar. Neglecting the pressure dropover the balance line 9 the pressure at the second port 92 is the sameas the pressure at the first port 91.

Optionally, one or more bypass lines can be disposed to limit the flowof process fluid through the different bearings 53, 54, 55, 43, 44, 45.In FIG. 5 a first bypass line 93 is shown, which is configured to bypassthe axial pump bearing 55, the torque converter 8, the drive unit 4, thesecond radial drive bearing 44 and the axial drive bearing 45. A firstthrottle 931 is disposed in the first bypass line 93 to regulate theflow of process fluid that passes through all the components that arearranged above the balance drum 7. The first throttle 931 can beconfigured e.g. as an orifice. Thus, a first part of the process fluidexiting the relief passage 73 flows through all the components 55, 8,53, 43, 4, 44, 45 arranged above the balance drum 7, and then via thefirst port 91 into the balance line 9, and a second part of the processfluid exiting the first relief passage 73 bypasses the components 55, 8,53, 43, 4, 44, 45 and directly enters the balance line 9. In FIG. 5, thefirst bypass line 93 is shown as an external line. The entrance to thefirst bypass line 93 is located at the common housing 2 at a locationbetween the first balance drum 7 and the axial pump bearing 55(regarding the axial direction A). From the entrance the first bypassline 93 extends towards the balance line 9 and opens out into thebalance line 9. However, it is also possible and for many applicationseven preferred, that the first bypass line 93 is an internal line, whichis completely located inside the common housing 2. For this purpose, thefirst bypass line 93 can constitute a direct flow communication betweenthe back side 72 and the first port 91, or the volume above the axialdrive bearing 45, respectively. Configuring the first bypass line 93 asan internal line has the advantage that the number of openings requiredat the common housing 2 can be reduced.

Optionally, a second bypass line 94 can be provided, which is configuredto bypass the second radial pump bearing 54 at the non-drive end 52 ofthe pump shaft 5. A second throttle 941 is disposed in the second bypassline 94 to regulate the flow of process fluid that passes through thesecond radial pump bearing 54. The second throttle 941 can be configurede.g. as an orifice. Thus, a first part of the process fluid flowingthrough the balance line 9 flows through the second radial pump bearing54 to the low pressure side, and a second part of the process fluidflowing through the balance line 9 bypasses the second radial pumpbearing 54 and directly enters the low pressure side. In FIG. 5, thesecond bypass line 94 is shown as an external line connecting thebalance line 9 with the low pressure side adjacent to the pump inlet 21.The entrance to the second bypass line 94 is located at the balance line9. From there the second bypass line 94 extends towards the commonhousing 2 and is connected to an opening at the common housing, whichopening is arranged at a location, where essentially the low pressureprevails. However, it is also possible and for many applications evenpreferred, that the second bypass line 94 is an internal line, which iscompletely located inside the common housing 2. For this purpose, thesecond bypass line 94 can be configured to constitute a direct flowcommunication between the second port 92 or the volume below the secondradial pump bearing 54, respectively, and the low pressure side or thepump inlet 21, wherein the flow communication bypasses the second radialpump bearing 54. Configuring the second bypass line 94 as an internalline has the advantage that the number of openings required at thecommon housing 2 can be reduced.

Reverting to the numerical example that has been given with reference tothe first embodiment of the pump, the following different pressures canprevail at and in the second embodiment of the pump 1: When, as anexample, the pump 1 is deployed at the sea ground in a depth of 250 mbelow the water surface, the low pressure prevailing at the pump inlet21 is e.g. 25 bar. The pump 1 can increase the pressure by 175 bar.Thus, the high pressure at the high pressure outlet 22 is 200 bar.Taking into consideration that there is also a pressure drop between theback side 72 and the first port 91 as well as over the second radialpump bearing 54, the pressure drop over the balance drum 7 is less thanthe pressure increase generated by the pump 1, i.e. the differencebetween the high pressure and the low pressure. For example, thepressure drop over the balance drum 7 can be 160 bar, the pressure dropover the bearings 55, 53, 43, 44, 45, the torque converter 8 and thedrive unit 4 can be 10 bar and the pressure drop over the second radialpump bearing 54 can be 5 bar. Accordingly, the intermediated pressureprevailing at the back side 72 is about 40 bar. The pressure at thefirst port 91, the second port 92 and within the balance line 9, isapproximately 30 bar (neglecting the pressure drop over the balance line9).

In other configurations the pump unit 3 can comprise a first pumpsection having a first set of impellers 31 and a second pump sectionhaving a second set of impellers 31.

The first pump section comprising the first set of impellers 31 and thesecond pump section comprising the second set of impellers 31 can bearranged in an inline arrangement or in a back-to-back arrangement.

In an inline arrangement the first set of impellers 31 and the secondset of impellers 31 are configured such that the axial thrust generatedby the action of the rotating first set of impellers 31 is directed inthe same direction as the axial thrust generated by the action of therotating second set of impellers 31. Thus, the flow of process fluid,which is generated by the second set of impellers 31, is directed in thesame direction as the flow of process fluid, which is generated by thefirst set of impellers 31. In such an arrangement the common housing 2can include an additional inlet and with an additional outlet, such thatthe first pump section and the second pump section can be used as twopumps.

In a back-to-back arrangement the first set of impellers 31 and thesecond set of impellers 31 are configured such that the axial thrustgenerated by the action of the rotating first set of impellers 31 isdirected in the opposite direction as the axial thrust generated by theaction of the rotating second set of impellers 31. Thus, the flow ofprocess fluid, which is generated by the second set of impellers 31, isdirected in the opposite direction as the flow of process fluid, whichis generated by the first set of impellers 31. In such an arrangementthe common housing 2 can be provided with an additional inlet and withan additional outlet, such that the first set of impellers 31 and thesecond set of impellers 31 can be used as a first pump section and asecond pump section. The first pump section and the second pump sectioncan be used as two pumps.

For many applications the back-to-back arrangement is preferred becausethe axial thrust acting on the pump shaft 5, which is generated by thefirst set of impellers 31 counteracts the axial thrust, which isgenerated by the second set of impellers 31. Thus, the two axial thrustscompensate each other at least partially.

What is claimed:
 1. A centrifugal pump for conveying a process fluid,comprising: a pump unit; a drive unit configured to drive the pump unit;a pump inlet configured to receive the process fluid; a pump outletconfigured to discharge the process fluid, the pump unit comprising atleast one impeller configured to convey the process fluid from the pumpinlet to the pump outlet, and a pump shaft, on which each impeller ismounted, the drive unit comprising a drive shaft configured to drive thepump shaft, and an electric motor configured to rotate the drive shaftabout an axial direction; a plurality of bearings configured to supportthe pump shaft and the drive shaft; and a hydrodynamic coupling having acasing and configured to hydrodynamically couple the drive shaft to thepump shaft by a transmission fluid, at least one of the plurality ofbearings is arranged in the casing of the hydrodynamic coupling.
 2. Thepump in accordance with claim 1, the hydrodynamic coupling comprises apump wheel connected to the drive shaft, a turbine wheel connected tothe pump shaft, and a stator arranged between the pump wheel and theturbine wheel configured to guide the transmission fluid.
 3. The pump inaccordance with claim 1, wherein the plurality of bearings comprises afirst radial pump bearing configured to support the pump shaft, thefirst radial pump bearing is arranged between the pump unit and a driveend of the pump shaft, and the first radial pump bearing is arranged inthe casing of the hydrodynamic coupling.
 4. The pump in accordance withclaim 1, wherein the plurality of bearings comprises a first radialdrive bearing configured to support the drive shaft, the first radialdrive bearing is arranged between the pump unit and the electric motor,and the first radial drive bearing is arranged in the casing of thehydrodynamic coupling.
 5. The pump in accordance with claim 1, whereinthe pump is a process fluid lubricated pump, and includes a commonhousing, the pump unit and the drive unit are arranged in the commonhousing, and the plurality of bearings is configured to receive theprocess fluid as lubricant and coolant.
 6. The pump in accordance withclaim 1, wherein the hydrodynamic coupling is configured to receive theprocess fluid as the transmission fluid.
 7. The pump in accordance withclaim 1, wherein the pump is a seal-less pump without a mechanical seal.8. The pump in accordance with claim 1, further comprising a balanceline configured to recirculate the process fluid from a high pressureside to a low pressure side of the pump.
 9. The pump in accordance withclaim 1, wherein the process fluid is configured to pass through andcool the electric motor.
 10. The pump in accordance with claim 8,wherein the balance line is arranged and configured to receive theprocess fluid discharged from the drive unit.
 11. The pump in accordancewith claim 5, further comprising an external cooling loop configured tocool and lubricate the drive unit and the plurality of bearings, theexternal cooling loop comprising a heat exchanger configured to cool theprocess fluid, the heat exchanger is arranged outside the common housingand configured to receive the process fluid from the common housing andto recirculate the process fluid to the common housing.
 12. The pump inaccordance with claim 1, wherein the pump unit comprises an intermediatetake-off connected to a cooling loop, the intermediate take-offconfigured to supply the process fluid to the cooling loop with apressure that is larger than a pressure of the process fluid at the pumpinlet, and the cooling loop is configured to supply process fluid to atleast one bearing of the plurality of bearings or to the drive unit. 13.The pump in accordance with claim 1, further comprising at least onebalance drum or center bush or throttle bush fixedly connected to thepump shaft and defining a front side facing the pump unit and a backside, and further comprising a relief passage, between the balance drumand a stationary part stationary with respect to the common housing, therelief passage is extending in axial direction along the balance drumfrom the front side to the back side.
 14. The pump in accordance withclaim 1, wherein the pump is a vertical pump with the pump shaftextending in the direction of gravity, and the drive unit is arranged ontop of the pump unit.
 15. The pump in accordance with claim 1, whereinthe pump is configured to be installed on a sea ground.
 16. The pump inaccordance with claim 15, wherein the pump is a water injection pumpconfigured to inject seawater into a subterranean region